As is known a compressor element of a screw compressor contains a housing with two meshing rotors, of which the helical parts are called the rotor bodies.
As is known, one of the rotors is constructed in the form of a male rotor with lobes, while the other rotor is constructed in the form of a female rotor with grooves, in which the lobes of the male rotor mesh in a known way.
Each rotor has a shaft whereby the rotor is supported in the housing by means of axial and radial bearings.
Generally one of these rotors is driven by a motor via a drive gearwheel on the shaft of the rotor, and in turn this rotor drives the other rotor, via a gearwheel transmission formed by gearwheels on the shafts of both rotors, or otherwise.
The housing of the screw compressor element has an inlet close to one end of the rotors through which a gas to be compressed is drawn in (i.e. the inlet side of the compressor element) during the operation of the screw compressor, and this screw compressor element has an outlet close to the other end of the rotors through which compressed gas is expelled (i.e. the ‘outlet side’ of the compressor element).
As is known, when driving the rotors by means of the aforementioned motor, a gas or mixture of gases to be compressed, such as air, is drawn in at the aforementioned inlet by the meshing of the rotors and then compressed between the two rotors and finally driven out from the outlet side of the compressor element at a certain outlet pressure.
An axial gas force will thereby be exerted on the rotors, which during nominal operating conditions is directed from the outlet side to the inlet side, as the gas pressure on this outlet side is higher than the gas pressure on the inlet side. When starting up, this gas force is absent as at that time no gas is being compressed and the gas pressures on the inlet side and the outlet side are practically equal.
During operation of the screw compressor element, the rotors also experience axial forces that are exerted by the aforementioned motor on the drive gearwheel of one of the two rotors, such as is the case for example with an oblique or helical toothing of the drive gearwheel, and/or when there is a gearwheel transmission between the rotors, due to the axial forces attributable to the torque transmission between the gearwheels of this gearwheel transmission.
The resulting axial forces that act on the rotors can change direction during operation of the screw compressor, such that during starting up the forces exerted by the gearwheels, or more generally by the drive, can have the upper hand, whereas during nominal operation of the screw compressor element, the gas forces are generally more pronounced and determine the direction of the resulting axial forces.
Due to this reversal of the direction of the resulting axial forces, it is consequently necessary for the axial forces to be able to be accommodated in each direction, in order to immobilise the rotors in the housing as much as possible, on the one hand in order to avoid undesired leakage losses arising through the axial head clearance between the head surface of the rotors, and the corresponding head surface of the outlet side of the housing being too large, and on the other hand to avoid one of the head surfaces of the rotors coming into contact with the housing, which would lead to undesired friction and wear.
From EP 0.867.628 applications are also known in which the head surface clearance is kept constant, whereby the clearance is measured by a sensor and the rotor can be moved in the axial direction by means of a dual-action magnetic actuator that moves the rotor in the one or the other direction when the clearance deviates from the set desired value.
Such applications are rather complex and expensive.
Other applications are also known in which the rotor is kept in place in the axial direction by applying dual-action axial bearings, for example four-point contact bearings, that can accommodate the forces on the rotor in both directions.
However these dual-action axial bearings have an inherent axial play. As a result, when the resultant forces on the rotors reverse, the rotors can move over a distance which, neglecting the deformation of the bearings, is equal to this play.
When designing the compressor element this must be taken into account by providing a sufficiently large head clearance to avoid the rotor making undesired contact with the housing, which would then mean a loss of efficiency as a result of leakage losses.
Dual-action bearings also generate relatively large mechanical losses and churning losses as a result of the lubricants in the bearing being churned.
On the other hand, a certain inherent clearance in such dual-action bearings is necessary because if the clearance is too small, the risk of a three-point contact increases, which is undesired as in this case the risk of slipping in the bearing increases substantially, and consequently either excessive lubricant has to be applied, or the operating range of the compressor has to be limited, such that low loads at high speeds in particular are to be avoided.
Such compensating measures are already known in the form of a “balancing piston”, as described in BE 1.013.221 for example, whereby a hydraulic or pneumatic piston is used to exert an axial force on the rotor concerned in order to counteract the gas forces acting on the rotor and thus to relieve the main bearing, such that a smaller bearing can be selected. The balancing piston is connected to the outlet of the compressor element, so that the force exerted by the balancing piston is proportional to the outlet pressure.
A disadvantage is the cost of such a balancing piston. Another disadvantage is that the main bearing always experiences the full load that is due to the resultant of the gas forces and the forces developed on the drive gearwheel and/or the gearwheel transmission between the rotors, such that the main bearing has to be designed for this.
One way to avoid the load on the main bearing changing direction is to provide a spring that exerts an elastic force on the rotor in the axial direction, in the opposite direction to the resulting forces on the gearwheels of the rotor concerned and in the same direction as the gas forces exerted on the rotor.
If this axial elastic force is sufficiently large to fully offset the resulting forces on the gearwheels of the rotor when starting up the compressor element, there will be no reversal of the resultant of the forces on the rotor, such that in this case a single-action axial bearing can suffice as the main bearing.
The aforementioned spring is affixed between the housing and the rotor concerned, and consequently requires the introduction of a supplementary axial bearing to transmit the axial elastic force between the stationary housing and the turning rotor.
This supplementary bearing has the disadvantage that it increases the cost and also brings about additional losses.
Although the spring prevents the direction of the load on the main bearing reversing during the start of the compressor element, this same spring will detrimentally increase the load on the main bearing during nominal operation as the resulting forces on the gearwheels of the rotor are at least partially cancelled out by the spring, and these resulting forces consequently can no longer counteract the gas forces during nominal operation to the same extent as in a situation without a spring.